Rotating-Plate Radial Turbine in Gas-Turbine-Cycle Configurations

ABSTRACT

A novel power-producing concept is disclosed, employing a rotating-plate radial gas turbine in various gas-turbine cycle configurations. The “rotating-plate radial gas turbine” is a rotating barrel with robust rectangular plates fitted into the turbine rotor, performing the function and the role of turbine blades, contained within a motionless rigid horizontal cylinder (casing). Combustion can take place in the spaces confined between adjacent rotating plates and the static cylinder, thus enabling a practical achievement of the Atkinson cycle (constant-volume heat addition) with improved cycle thermal efficiency. Alternatively, two or more compressor stages can be combined to feed a single rotating-plate radial gas turbine in cascades, thus gradually increasing pressure of working gas within a volume bordered by adjacent un-cooled rotating plates of the radial gas turbine and the casing. Alternatively, a single compressor may be combined with one or more stages of an axial turbine for cascade feeding of a single rotating-plate radial gas turbine. This “isochoric stuffing” effect enables achievement of significantly and even drastically improved gas-turbine-cycle thermal efficiencies. Cycle heat addition may be either isobaric or isochoric in either an open-cycle or a closed-cycle configuration. Using a sufficiently efficient radial gas turbine, it is theoretically possible to obtain 100% cycle thermal efficiency in a simple radial-gas-turbine configuration with appropriately chosen compressor-stages compression pressure ratios.

BACKGROUND OF THE INVENTION

The concept disclosed in this invention relates to a means or a device (gas-turbine power plant) and method for generating of power from heat by employing an improved component—rotating-plate radial gas turbine. The concept has a potential to enable achievement of significantly and even drastically improved gas-turbine-cycle thermal efficiencies.

In today's conditions of an ongoing and growing energy crisis and recently increased and accelerating global climate changes and global warming phenomena, world's governments, large companies and engineers/scientists are intensely interested in developing more energy-efficient and improved power-producing systems. Increased concentration of CO₂-gas in the atmosphere is today recognized as the main culprit of global warming phenomena. Together with water vapor (H₂O), CO₂ represents the main and an unavoidable ingredient of flue gases resulting from combustion of fossil fuels (coal, fuel oil/petroleum, natural gas) needed for main industrial sectors of any country: power generation, transportation, chemical/petrochemical industry and metallurgy.

One of the simplest and the most logical way of reducing environmental pollution and mitigating the energy crisis is to improve thermal (energy) efficiency of the thermal-to-mechanical (or electrical) energy conversion. Typically used steam- and gas-turbine power plants have cycle thermal efficiencies between 35% and 45% at LHV (lower heating value of fuel). The most efficient gas turbines have cycle thermal efficiency of ˜42%. The most efficient steam-turbine power-plants at ultra-super-critical steam parameters have cycle thermal efficiencies nearing -50%. However, with steam-boiler losses, the overall maximum steam-turbine cycle thermal efficiency is ˜45%. Steam turbines for nuclear power plants have even lower cycle thermal efficiencies, ˜34-44%. Energy efficiency situation is particularly unenviable in the transportation sector, where gas turbines are typically used for propelling of aircrafts and internal-combustion reciprocating engines are typically used for propelling of motor vehicles. Typically somewhat higher cycle thermal efficiencies of internal-combustion reciprocating engines (compared to gas turbines) are often diminished and offset by significantly increased engine weights for identical power outputs. In addition, internal-combustion reciprocating engines are much more complex in their design/construction.

Nowadays, combined-cycle gas & steam turbine plants have the highest cycle thermal efficiency (nearing 60% at LHV) of all thermal engines. In addition, new hybrid concepts have been developed with small-size gas turbines combined with fuel cells, with claimed overall thermal efficiency near 70%. However, this advanced power-producing concept can be applied in small sizes only, for so-called distributed power generation. Both above mentioned advanced power-producing concepts use primarily natural gas (or other hydrocarbon fuels) as fuel and are therefore dependent on its/their availability. An alternative fuel, synthetic natural gas (SNG or “syngas”) can be obtained by the coal gasification, since coal is the most abundant fossil fuel on the Earth.

Although modern power-producing systems achieve ever-increasing cycle thermal efficiencies, it appears that these improvements are not sufficient to enable converting bulk of Earth's fossil-fuels reserves into mechanical/electrical energy in an efficient, enough technically simple and enough cheap manner. With current trend, future of our planet seems to be doomed, unfortunately. People on planet Earth need even more energy-efficient and radically improved, technically simpler and even more reliable power-producing systems and/or methods to satisfy their needs for energy and yet to preserve health of our environment for generations to come.

BRIEF SUMMARY OF THE INVENTION

An object of the invention is to provide an improved gas-turbine power plant component, a rotating-plate radial gas turbine, in the form of a rotating barrel with robust rectangular plates fitted into a radial-turbine rotor, contained within a motionless rigid horizontal cylinder (casing).

Another object of the invention is to provide means/device (gas-turbine power plant) for generating of power from heat by employing the improved component—said rotating-plate radial gas turbine using either isobaric (constant-pressure) or isochoric (constant-volume) combustion chamber.

Yet another object of the invention is to provide an improved method of using said rotating-plate radial gas turbine by employing two or more compressor stages to jointly feed the single said rotating-plate radial gas turbine in cascades from the radial direction, thus gradually increasing working-gas pressure across the volume of the said radial gas turbine for achievement of a greater cycle output and a higher cycle thermal efficiency.

Yet another object of the invention is to provide an improved method of using said rotating-plate radial gas turbine by employing two or more axial turbine stages and a single-stage compressor to jointly feed the single said rotating-plate radial gas turbine in cascades from the radial direction, thus gradually increasing working-gas pressure across the volume of the said radial-gas-turbine for achievement of a greater cycle output and a higher cycle thermal efficiency.

Still another object of the invention is to provide an improved method of using said rotating-plate radial gas turbine by employing constant-volume (isochoric) combustion of gaseous or liquid fuel in the spaces confined between adjacent rotating plates of the said rotating-plate radial gas turbine and its casing, enabling a practical approach to achievement of the Atkinson cycle. The said method can be used in any of the said rotating-plate radial gas-turbine configurations: with a single-stage compressor, with two or more compressor stages, or with two or more axial turbine stages and a single-stage compressor.

Still another object of the invention is to provide an improved method of using said rotating-plate radial gas turbine by employing two or more compressor stages to jointly feed the single said rotating-plate radial gas turbine in cascades from the radial direction (gradually increasing working-gas pressure across the volume of the said radial gas turbine) without any heat input/addition, thus potentially achieving the highest possible cycle thermal efficiency of 100%.

Still another object of the invention is to highlight the fact that the said rotating-plate radial gas turbine and any of the said improved methods of using it can be applied also to corresponding closed-cycle configurations of the power-producing system, with constant-pressure (isobaric) heat input/addition. In addition, the closed-cycle configurations of the power-producing system without heat rejection/output would also potentially achieve the highest possible cycle thermal efficiency of 100%.

Still another object of the invention is to highlight the fact that the said rotating-plate radial gas turbine open-cycle configurations of the power-producing system and any of the said improved methods of using it can be applied also to aircraft engines and external-combustion engines in motor vehicles.

Finally, another object of the invention is to highlight the fact that the any of the said rotating-plate radial gas turbine closed-cycle configurations and any of the said improved methods of using it can be applied to a power-producing system using any fuel source, such as coal, nuclear fuel, renewable energy and waste energy (heat).

Provided it proves to be technically/technologically valid and feasible, this very efficient (in some special instances close to 100%) and conditionally zero-fuel-input power-generation concept would result in potentially huge cycle thermal efficiency improvements and hence likely in dramatic consequences in the future of energy supply & energy prices at the planet Earth. In addition, it could also become a very efficient means for reducing of greenhouse gases emissions and indirect establishing of their normal natural quantities in the Earth's atmosphere.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 depicts flow diagram of a simple open-cycle rotating-plate radial gas-turbine power-producing system with isobaric or isochoric combustion chamber.

FIG. 2 depicts flow diagram of an open-cycle rotating-plate radial gas-turbine power-producing system with isobaric or isochoric combustion chamber and with a two-stage intercooled compressor.

FIG. 3 depicts flow diagram of an open-cycle rotating-plate radial gas-turbine power-producing system with isobaric or isochoric combustion chamber and with a pre-cooled compressor using an absorption refrigeration chiller.

FIG. 4 depicts flow diagram of an open-cycle rotating-plate radial gas-turbine power-producing system with isobaric or isochoric combustion chamber and with a two-stage pre-cooled/intercooled compressor using an absorption refrigeration chiller.

FIG. 5 depicts flow diagram of a simple open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages and with isobaric or isochoric combustion chamber.

FIGS. 6 & 7 depict two flow diagrams of a simple closed-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages, with isobaric heat input/addition (at two different locations) and without any heat rejection.

FIG. 8 depicts flow diagram of a simple open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages and without any heat input.

FIG. 9 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages, with isobaric or isochoric combustion chamber and with a recuperator between the first two compressor stages.

FIG. 10 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages, with isobaric or isochoric combustion chamber and with a 3-pressure-levels recuperator.

FIG. 11 depicts flow diagram of an open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages, with isobaric or isochoric combustion chamber and with the first compressor stage pre-cooled using an absorption refrigeration chiller.

FIG. 12 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with 3 compressor stages, with isobaric or isochoric combustion chamber, with a recuperator between first two compressor stages and with the first compressor stage pre-cooled using an absorption refrigeration chiller.

FIG. 13 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages, with isobaric or isochoric combustion chamber, with a 3-pressure levels recuperator and with pre-cooled compressor stages using an absorption refrigeration chiller.

FIG. 14 depicts flow diagram of a simple open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages and with isobaric or isochoric combustion chamber.

FIG. 15 depicts flow diagram of a simple closed-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages, with isobaric heat input/addition and with isobaric heat output/rejection.

FIG. 16 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages, with isobaric or isochoric combustion chamber and with a typical recuperator.

FIG. 17 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages and two compressor stages, with isobaric or isochoric combustion chamber and with a recuperator located between the two compressor stages.

FIG. 18 depicts flow diagram of an open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages, with isobaric or isochoric combustion chamber and with a pre-cooled compressor using an absorption refrigeration chiller.

FIG. 19 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages, with isobaric or isochoric combustion chamber, with a typical recuperator and with a pre-cooled compressor using an absorption refrigeration chiller.

FIG. 20 depicts flow diagram of a recuperated open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with two axial turbine stages and two compressor stages, with isobaric or isochoric combustion chamber, with a recuperator located between the two compressor stages and with the first compressor stage pre-cooled using an absorption refrigeration chiller.

FIGS. 21 and 22 depict thermodynamic temperature-entropy diagrams corresponding to a simple open-cycle power-producing system, with isobaric combustion chamber (max. cycle temperature 1500 K), using a three-pressures-fed and a five-pressures-fed rotating-plate radial gas turbine, respectively, with three and five compressor stages, respectively.

FIG. 23 depicts thermodynamic temperature-entropy diagram corresponding to a simple open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine, with isobaric combustion chamber (max. cycle temperature 1500 K) and with two axial turbine stages and one compressor stage.

FIG. 24 depicts thermodynamic temperature-entropy diagram corresponding to a simple open-cycle power-producing system using a three-pressures fed rotating-plate radial gas turbine with three compressor stages and without any heat input.

DETAILED DESCRIPTION OF THE INVENTION

In this chapter a detailed description of preferred rotating-plate radial gas-urbine cycle configurations and explanation of corresponding thermodynamic diagrams will be given, with a more detailed explanation of the rotating-plate radial gas-turbine working principle.

This invention advocates the use of an innovation in the gas-turbine power-plants—a rotating-plate radial gas turbine, which is expected to be a significant improvement and to dramatically increase the GT power-plant cycle thermal efficiency. As already mentioned previously, the rotating-plate radial gas turbine is planned in the form of a rotating barrel with robust rectangular plates (“rotating plates”), performing the function and the role of turbine blades, fitted into a radial-turbine rotor, contained within a motionless rigid horizontal cylinder (casing). Physically it very much resembles an ordinary water mill with its simple and robust design and is similar to an impulse water turbine (“Pelton turbine”). Also, it is physically very similar to rotating-plate regenerative air pre-heaters typically used for preheating of combustion air in conventional steam boilers. However, its operating principle is identical to that of radial reaction water turbines (“Francis turbine”), where pressure decreases across the water turbine (to atmospheric pressure) and transfers its energy to the rotating wheel. In addition and for better explanation, the operating principle of a rotating-plate radial gas turbine is almost identical also to that of a positive-displacement rotary vane pump that consists of vanes mounted to a rotor that rotates inside of a larger circular cavity. Generally, a positive displacement pump causes a fluid (liquid) to move by trapping a fixed amount of it then forcing (displacing) that trapped volume into the discharge pipe.

The main difference between the radial gas turbine and water turbines/mills and positive-displacement rotary vane pumps is, of course, its working fluid: a compressible fluid, that is,. a gas (air in open-cycle configurations). This fact enables that compressible-fluid phenomena, such as isochoric (constant-volume) combustion, can be considered and employed in the rotating-plate radial turbine design, as explained later. Working gas is fed and exhausted from the radial direction of the rotating-plate radial-gas-turbine, thru corresponding side/top/bottom openings in the said casing (static cylinder). In general, the rotating-plate radial-turbine cycle heat addition may be either isobaric or isochoric in either an open-cycle or a closed-cycle configuration.

When used in a simple gas-turbine configuration, the rotating-plate radial turbine can increase cycle thermal efficiency by eliminating the ever-present loss in ordinary axial gas-turbines: fraction of compressed working gas (typically air in open-cycle configurations) needed for cooling of the rotor and vanes of an ordinary axial gas turbine. The robust and simple rectangular-shaped rotating blades of the proposed radial gas turbine do not need to be cooled at all. The same fact enables also use of potentially higher gas-turbine inlet temperatures and hence higher GT power outputs and GT cycle thermal efficiencies. It may also be obvious to an ordinary person skilled in the art that the proposed power-producing concept can readily be employed in motor-vehicle engines and aircraft engines.

In addition, this invention proposes a further improvement related to the use of the rotating-plate radial gas turbine: isochoric (constant-volume) combustion. Instead of typically used isobaric (constant-pressure) combustion chamber, an isochoric (constant-volume) combustion “chamber” can be used. Actually, isochoric combustion would take place in the spaces confined between adjacent rotating plates of the rotating-plate radial gas turbine and its casing, where also a gaseous or a liquid fuel has to be injected and ignited by means of an electric spark. Thus, a practical gas-turbine achievement of the Atkinson cycle (adiabatic compression and expansion, isochoric heat addition and isobaric heat rejection) is enabled, with improved cycle thermal efficiency. Also, cycle power output would be increased compared to the output of conventional internal-combustion piston engines employing the same Atkinson cycle. The thermal efficiency and power output improvements of the Atkinson cycle are the result of an increased expansion pressure ratio (equivalent to the ratio of the maximum combustion temperature and the compressor discharge temperature) for the same compression ratio of the Brayton cycle (isobaric heat addition/rejection GT cycle) or the Otto cycle (isochoric heat addition/rejection piston engine cycle). Not to be underestimated is also the fact that such a rotating-plate radial GT external-combustion engine would be much lighter, smaller and construction-wise simpler compared to an internal-combustion piston engine of the same power output.

Perhaps the most innovative and promising in the proposed concept of the rotating-plate radial gas turbine is the method of multi-pressure radial-turbine gas feeding from consecutive compressor or axial-turbine stages thru corresponding side/top/bottom openings in the rotating-plate radial-gas-turbine casing. Two or more compressor stages (of any type) can be combined together to jointly feed a single said rotating-plate radial gas turbine in cascades from the radial direction, thus gradually increasing pressure of the air/working gas within a constant volume bordered by adjacent un-cooled rotating plates of the said radial gas turbine and the said casing. Alternatively, a single compressor may be combined with one or more axial turbine stages for cascade feeding of a single rotating-plate radial gas turbine. Such a gradual pressure increase is a result of the “isochoric stuffing” effect (mixing of working gas at different pressure levels in a confined and constant volume), which enables achievement of significantly and even drastically improved gas-turbine-cycle thermal efficiencies and greater power outputs.

According to this invention, in some special instances, with appropriately and wisely chosen compression pressure ratios of consecutive compressor stages using a sufficiently efficient rotating-plate radial gas turbine, it is theoretically possible to obtain nearly 100% cycle thermal efficiency in a simple closed-cycle or open-cycle radial gas-turbine configuration. What is even more appealing is that such a simple open-cycle radial gas-turbine configuration would (theoretically for now) be able to extract some work from the surrounding environmental air without any heat input.

For illustration and better explanation, the following ideal-gas-model thermodynamic analysis is given below in support of statements in previous paragraphs.

Ideal Gas Model—Realistic Flow Situation—Isochoric (constant-volume) gas entry into a confined volume “V” containing initial mass of gas m₀ at initial pressure p₀ and temperature T₀, as follows:

The average gas-mixture temperature (T_(mix,1)) can be estimated from the mixing rule of two gas quantities (m₀ and m₁), as follows:

$T_{{mix},1} = \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}}} \right)}{\left( {m_{0} + m_{1}} \right)}$

The average gas-mixture pressure (p_(mix,1)) can be estimated from the perfect-gas equation and the above expression for the average gas-mixture temperature, as follows:

$\left. \begin{matrix} {\frac{V}{R_{gas}} = {{const}.}} \\ {= \frac{m_{0} \cdot T_{0}}{p_{0}}} \\ {= \frac{\left( {m_{0} + m_{1}} \right) \cdot T_{{mix},1}}{p_{{mix},1}}} \\ {= \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}}} \right)}{p_{{mix},1}}} \end{matrix}\Rightarrow p_{{mix},1} \right. = {p_{0} \cdot \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}}} \right)}{T_{0} \cdot m_{0}}}$

The maximum gas quantity (m_(1,max)) that can entry the volume V at pressure p_(mix,1)=p₁ is as follows:

$\left. \Rightarrow m_{1,\max} \right. = {\left. {m_{0} \cdot \frac{T_{0}}{T_{1}} \cdot \left( {\frac{p_{1}}{p_{0}} - 1} \right)}\Rightarrow\frac{{\overset{.}{m}}_{1,\max}}{{\overset{.}{m}}_{0}} \right. = {\frac{m_{1,\max}}{m_{0}} = {\frac{T_{0}}{T_{1}} \cdot \left( {\frac{p_{1}}{p_{0}} - 1} \right)}}}$ (in  time  Δ t)

In terms of gas mass flow rate, that is, the mass of gas flowing in time “Δt”, the initial mass of gas m₀ at initial gas parameters can be expressed as a “virtual initial mass flow rate” (Δm₀/Δt) knowing the gas flow rate of the first compressor stage (Δm_(1,max)/Δt), as follows:

$\left. \Rightarrow\frac{\Delta \; m_{0}}{\Delta \; t} \right. = {\left. {\frac{\Delta \; m_{1,\max}}{\Delta \; t} \cdot \frac{T_{1}}{T_{0}} \cdot \frac{1}{\left( {\frac{p_{1}}{p_{0}} - 1} \right)}}\Rightarrow{\overset{.}{m}}_{0} \right. = {{\overset{.}{m}}_{1,\max} \cdot \frac{T_{1}}{T_{0}} \cdot \frac{1}{\left( {\frac{p_{1}}{p_{0}} - 1} \right)}}}$ (in  time  Δ t)

Considering initial mass/pressure/temperature of gas (m₀+m₁, p_(mix,1) and T_(mix,1)) in the volume V, the flow situation becomes as follows after entry of the same gas at a different (higher) pressure (p₂):

The average gas-mixture temperature (T_(mix,2)) can be estimated from the mixing rule of three gas quantities (m₀, m₁ and m₂), as follows:

$T_{{mix},2} = \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}} + {T_{2} \cdot m_{2}}} \right)}{\left( {m_{0} + m_{1} + m_{2}} \right)}$

The average gas-mixture pressure (p_(mix,2)) can be estimated from the perfect-gas equation and the previous and the above expressions for the average gas-mixture temperatures, as follows:

$\begin{matrix} {\frac{V}{R_{gas}} = {{const}.}} \\ {= \frac{\left( {m_{0} \cdot m_{1}} \right) \cdot T_{{mix},1}}{p_{{mix},1}}} \\ {= \frac{\left( {m_{0} + m_{1} + m_{2}} \right) \cdot T_{{mix},2}}{p_{{mix},2}}} \\ {= \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}} + {T_{2} \cdot m_{2}}} \right)}{p_{{mix},2}}} \end{matrix}$ $\begin{matrix} {\left. \Rightarrow p_{{mix},2} \right. = {p_{{mox},1} \cdot \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}} + {T_{2} \cdot m_{2}}} \right)}{\left( {m_{0} + m_{1}} \right) \cdot T_{{mix},1}}}} \\ {= {p_{0} \cdot \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}} + {T_{2} \cdot m_{2}}} \right)}{T_{0} \cdot m_{0}}}} \end{matrix}$

The maximum gas quantity (m_(2,max)) that can entry the volume V at pressure p_(mix,2)=p₂ (that was previously at pressure p_(mix,1)=p₁) is as follows:

${\left. \Rightarrow m_{2,\max} \right. = {{\frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}}} \right)}{T_{2}} \cdot \left( {\frac{p_{2}}{p_{1}} - 1} \right)}\mspace{14mu} {or}}}\mspace{14mu}$ ${\overset{.}{m}}_{2,\max} = {{\frac{\left( {{T_{0} \cdot {\overset{.}{m}}_{0}} + {T_{1} \cdot {\overset{.}{m}}_{1}}} \right)}{T_{2}} \cdot \left( {\frac{p_{2}}{p_{1}} - 1} \right)}\mspace{14mu} \left( {{in}\mspace{14mu} {time}\mspace{14mu} \Delta \; t} \right)}$

The ratio of pressures p_(mix,2)=p₂ and p_(mix,1)=p₁ is given by the following expression:

$\begin{matrix} {\left. \Rightarrow\frac{p_{2}}{p_{1}} \right. = \frac{p_{{mix},2}}{p_{{mix},1}}} \\ {= \frac{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}} + {T_{2} \cdot m_{2}}} \right)}{\left( {{T_{0} \cdot m_{0}} + {T_{1} \cdot m_{1}}} \right)}} \\ {= \frac{\left( {{T_{0} \cdot {\overset{.}{m}}_{0}} + {T_{1} \cdot {\overset{.}{m}}_{1}} + {T_{2} \cdot {\overset{.}{m}}_{2}}} \right)}{\left( {{T_{0} \cdot {\overset{.}{m}}_{0}} + {T_{1} \cdot {\overset{.}{m}}_{1}}} \right)}} \end{matrix}$

Generally, for an arbitrary number “N” of multi-pressure fluid (gas) entries (from consecutive compressor or turbine stages) to a confined volume “V”, the following general expressions can be derived for the average gas-mixture pressure (p_(mix,N)) and the maximum possible gas quantity (m_(N,max)) that can entry the volume “V” at the pressure p_(mix,N), respectively, as follows:

$p_{{mix},N} = {{p_{0} \cdot \frac{\sum\limits_{j = 0}^{j = N}\left( {T_{j} \cdot m_{j}} \right)}{T_{0} \cdot m_{0}}} = {{p_{0} \cdot \frac{\sum\limits_{j = 0}^{j = N}\left( {T_{j} \cdot {\overset{.}{m}}_{j}} \right)}{T_{0} \cdot \overset{.}{m}}}\mspace{14mu} \left( {{in}\mspace{14mu} {time}\mspace{14mu} \Delta \; t} \right)}}$ $m_{N,\max} = {{\frac{\sum\limits_{j = 0}^{j = {N - 1}}\left( {T_{j} \cdot m_{j}} \right)}{T_{N}} \cdot \left( {\frac{p_{N}}{p_{N - 1}} - 1} \right)}\mspace{14mu} {or}}$ $\mspace{14mu} {{\overset{.}{m}}_{N,\max} = {{\frac{\sum\limits_{j = 0}^{j = {N - 1}}\left( {T_{j} \cdot {\overset{.}{m}}_{j}} \right)}{T_{N}} \cdot \left( {\frac{p_{N}}{p_{N - 1}} - 1} \right)}\mspace{14mu} \left( {{in}\mspace{14mu} {time}\mspace{14mu} \Delta \; t} \right)}}$

The above derived simple equations can be used to determine theoretical mass-flow-rate ratios in a multi-fed rotating-plate radial gas turbine, based on chosen compressor pressure ratios of compressor stages or chosen expansion pressure ratios of axial turbine stages. Hence, it is relatively straightforward to estimate resulting power outputs and cycle thermal efficiencies of various preferred configurations of the rotating-plate radial gas-urbine cycle.

For illustration and explanation purposes, the discharge temperature (T₂) of working gas at the compressor outlet depends on: the compressor inlet temperature (T₁), the chosen compression pressure ratio (CPR), the ratio of specific heats of the working gas (κ; the value is 1.40 for bi-atomic gases and air) and the compression isentropic efficiency (η_(i,comp)), as follows:

$T_{2} = {T_{1} \cdot {\left\lbrack {1 + {\left( {{C\; P\; R^{(\frac{K - 1}{K})}} - 1} \right) \cdot \frac{1}{\eta_{i,{comp}}}}} \right\rbrack \mspace{14mu}\lbrack K\rbrack}}$

Similarly, the exhaust-gas temperature (T₄) of working gas at the radial gas-turbine outlet depends on: the turbine inlet temperature (T₃), the available expansion pressure ratio (EPR), the ratio of specific heats of the working gas (κ; the value is 1.40 for bi-atomic gases and air) and the expansion isentropic efficiency (η_(i,exp)), as follows:

$T_{4} = {T_{3} \cdot {\left\lbrack {1 - {\left( {1 - {E\; P\; R^{(\frac{1 - K}{K})}}} \right) \cdot \eta_{i,\exp}}} \right\rbrack \mspace{14mu}\lbrack K\rbrack}}$

The following is an explanation and illustration of how speed of rotation of a rotary (axial or radial) turbine or compressor influences the machine size. If a rotating-plate radial gas turbine develops power P (W) with the mass flow rate m^((dot)) (kg/s) of working gas at a speed or rotation expressed in either revolutions per minute (revs/min) n, or as an angular velocity ω (rad/s or s⁻¹), it is possible to express the radial gas-turbine radius R in terms of these parameters, as follows:

$\left. \begin{matrix} {P = {F \cdot w}} \\ {= {F \cdot R \cdot \omega}} \\ {= {\overset{.}{m} \cdot w \cdot R \cdot \omega}} \\ {= {\overset{.}{m} \cdot R^{2} \cdot \omega^{2}}} \end{matrix}\Rightarrow\begin{matrix} {R = {\left( \frac{1}{\omega} \right) \cdot \sqrt{\frac{P}{\overset{.}{m}}}}} \\ {= {\left( \frac{60}{2 \cdot \pi \cdot n} \right) \cdot {\sqrt{\frac{P}{\overset{.}{m}}}\mspace{14mu}\lbrack m\rbrack}}} \end{matrix} \right.$

It can be noted that the gas-turbine radius R is inversely proportional to the turbine's speed or rotation expressed in either revolution per minute (revs/min) n, or as an angular velocity ω (rad/s or s⁻¹). Provided the GT diameter D is identical to the GT axial length L, it can be estimated from the GT volumetric or mass flow rate, V^((dot)) or m^((dot)), working gas density ρ and the GT speed of rotation expressed in either revolution per minute (revs/min) n, or as an angular velocity ω (rad/s or s⁻¹), as follows:

$\overset{.}{V} = {\frac{\overset{.}{m}}{\rho} = {{\frac{D^{2} \cdot \pi \cdot L}{4} \cdot \frac{n}{60}} = {\left. {\frac{D^{3} \cdot \pi}{240} \cdot n}\Rightarrow D \right. = {L = {\sqrt[3]{\frac{240 \cdot \overset{.}{m}}{\pi \cdot \rho \cdot n}} = {\sqrt[3]{\frac{8 \cdot \overset{.}{m}}{\rho \cdot \omega}}\mspace{14mu}\lbrack m\rbrack}}}}}}$

The above derived equation illustrates how enormously small could be diameters/lengths of gas turbines, axial or radial, at large speeds of rotation for small power levels (hence, small mass flow rates), such as those typically used in motor vehicles. Even aircraft engines, with their intermediate power levels, can be small enough at large turbine rotational speeds. The same conclusion is valid also for rotating-plate radial gas turbines, with an important remark that recommended rotational speeds in their case should not exceed ˜3000-3600 revs/min (rotational speed of electric generators), to ensure sufficient time for associated working-gas mixing processes within the volume bordered by the radial-turbine rotating plates and the motionless casing.

The first preferred configuration of the power-producing system using a rotating-plate radial gas turbine is a simple open-cycle configuration depicted in FIG. 1, comprising: a conventional compressor (1) (axial, radial, reciprocating or any suitable type) compressing air (working gas), an isobaric (constant-pressure) combustion chamber (4) fuelled by a gaseous or a liquid fuel, and a load, typically an electric generator (8), connected to said compressor (1) via a common rotating shaft, a rotating-plate radial gas turbine (7) (essentially a rotating barrel with robust rectangular plates fitted into the turbine rotor, performing the function and the role of turbine blades, contained within a motionless rigid horizontal cylinder or casing), connected to the said compressor (1) and the said electric generator (8) via the same said common rotating shaft. The said rotating-plate radial gas turbine (7) is fed with air (working gas) by the said compressor (1) from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing. It also exhausts the expanded combustion gas thru another side/top/bottom opening in the said cylinder/casing.

Instead of the said isobaric combustion chamber (4), an optional isochoric (constant-volume) heat addition can be employed in the spaces confined between adjacent un-cooled rotating plates and said static cylinder, where also a gaseous or a liquid fuel has to be injected and ignited by means of an electric spark, thus enabling a practical achievement of the Atkinson cycle (adiabatic compression and expansion, isochoric heat addition and isobaric heat rejection) with improved cycle thermal efficiency and greater cycle output.

The second preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 2, is similar to the one depicted in FIG. 1, wherein, in addition to said compressor (1), the gas-urbine configuration comprises also a second stage of the compressor (2), accompanied by an intercooler (17) between the two said compressor stages, for achievement of a greater cycle output and a higher cycle thermal efficiency.

The next preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 3, is similar to the one depicted in FIG. 1, wherein it additionally comprises an absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution. The said absorption refrigeration chiller (10) consists of: a generator (11) heated by the exhaust gas from the said radial gas turbine (7) and evaporating the refrigerant from the solution, a condenser (12) for bringing the refrigerant into the liquid state, an absorber (16) for absorbing cold refrigerant vapor into the liquid solution, both said condenser (12) and said absorber (16) rejecting heat to cooling water or air from the environment, an expansion valve (13) for bringing the refrigerant to a lower temperature and an evaporator (14) for evaporating of the refrigerant-by pre-cooling of air (working gas) at the inlet of the said compressor (1), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.

The next preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 4, is similar to the one depicted in FIG. 3, wherein, in addition to the said compressor (1) and the corresponding said evaporator (14), the gas-turbine configuration comprises also a second stage of the compressor (2), accompanied by an intercooler (17) and a corresponding second-stage evaporator (15) for pre-cooling of air (working gas) at the inlet of the said second-stage compressor (2). Both said intercooler (17) and said second-stage evaporator (15) are located between the two said compressor stages, for achievement of a greater cycle output and a higher cycle thermal efficiency.

Another preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 5, is similar to the one depicted in FIG. 1, wherein, in addition to said compressor (1), the gas-turbine configuration comprises also a second stage of the compressor (2), a third stage of the compressor (3) and, if need may be, a plurality of the compressor stages, combined in such a way as to feed said rotating-plate radial gas turbine (7) in cascades from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing. This way the pressure of air (working gas) within a constant volume bordered by adjacent un-cooled rotating plates of the said radial gas turbine (7) and the said casing is being gradually increased, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. Said isobaric or isochoric combustion chamber (4) is located between the highest-pressure stage of the said compressor and the said radial gas turbine (7).

FIGS. 21 and 22 depict corresponding thermodynamic temperature-entropy diagrams of the open-cycle configuration of FIG. 5 with isobaric combustion chamber (max. cycle temperature 1500 K), using a three-pressures-fed and a five-pressures-fed rotating-plate radial gas turbine, respectively, with three and five compressor stages, respectively. Assuming an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 2:1 for each compressor stage, this cycle configuration yields cycle thermal efficiency in the range between 45.7% (three compressor stages) and 54.6% (five compressor stages). It is to be noted that the real inlet temperature of the rotating-plate radial gas-turbine is much lower than the maximum cycle temperature (1500 K), ranging from ˜703 K (430° C.) (three compressor stages) to ˜843 K (570° C.) (five compressor stages). The reason for this is that heat is added only to the air-flow fraction flowing thru the highest-pressure compressor stage, whereas all different-pressure-levels air-flow fractions are being mixed between the rotating plates of the radial gas turbine (7). These significantly lower radial GT inlet temperatures are one important and additional reason for omitting of cooling of the rotating plates (radial gas turbine “blades”). The main reason for the rotating plates to remain un-cooled is their robust and simple construction—they are just rectangular steel plates.

Logically, a closed-cycle gas-turbine power-producing system similar to the open-cycle multi-pressures-fed radial gas-turbine power-producing system depicted in FIG. 5 is also possible to configure, wherein a'suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture is being circulated. Instead of said isobaric/isochoric combustion chamber (4), such a closed-cycle gas-turbine configuration would comprise a closed-type gas heater (4″), located between the outlet of the highest-pressure stage of said compressor and the inlet of said radial gas turbine (7), which can use any fuel source, such as fossil fuels, nuclear fuels, renewable energy sources or waste heat. In addition, the configuration would have to include a closed-type heat rejection device, typically a heat exchanger cooled by environmental water or air. Such a closed-cycle radial GT configuration would be suitable for coal-fired and nuclear power plants of any type. For example, assuming helium as a working gas, maximum cycle temperature of 1123 K (850° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, this cycle configuration yields cycle thermal efficiency ranging from ˜41.2% (three-pressures-fed radial GT) to ˜68.8%-(five-pressures-fed radial GT).

The most interesting and the most preferred of the proposed open-cycle radial GT configurations is an open-cycle gas-turbine power-producing system depicted in FIG. 8, similar to the one depicted in FIG. 5, but with an essential difference: omission of said isobaric/isochoric combustion chamber (4) or any other heat addition at all. Such an open-cycle multi-pressures-fed radial GT configuration would employ solely the said effect of a gradual increase of the air (working-gas) pressure within a constant volume bordered by adjacent rotating plates of said radial gas turbine (7) and the said static casing as a result of feeding the said rotating-plate radial gas turbine (7) in cascades from the radial direction by combining multiple stages of the said compressor.

With respect to said, FIG. 24 depicts thermodynamic temperature-entropy diagram corresponding to such a simple open-cycle power-producing system using a three-pressures-fed rotating-plate radial gas turbine with three compressor stages and without any heat input. Assuming an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, the most preferred radial GT configuration could achieve a positive cycle output (from outside air; no heat addition!) and the maximum possible cycle thermal efficiency of 100%. It is to be noted that the radial GT inlet temperature (382 K) is lower than the highest-pressure compressor-stage discharge temperature (421 K). Consequently, the radial GT exhaust temperature (284.5 K) is also slightly lower than the ambient temperature (standard 288 K or 15° C.), which indicates that the cycle produces some positive power output at the expense of the energy from the outside air (cooling of outside air), again without any heat addition.

Similarly, the most preferred closed-cycle radial GT configuration is a closed-cycle gas-turbine power-producing system depicted in FIGS. 6 and 7, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, similar to the one depicted in FIG. 8, but with an essential difference: omission of any heat rejection device, wherein the said closed-type gas heater (4″) is located: (a) either between the outlet of the highest-pressure stage of said compressor and the inlet of said radial gas turbine (7) (for higher-temperature heat addition), or (b) between the outlet of the said radial gas turbine (7) and the inlet of the lowest-pressure stage of the said compressor (for lower-temperature heat addition). Similarly as with the same open-cycle radial GT configuration, assuming an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, the most preferred radial GT configuration could achieve a positive cycle output and the maximum possible cycle thermal efficiency of 100%. Such a closed-cycle radial GT configuration would be the most suitable and preferred for coal-fired and nuclear power plants of any type.

Another preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine, depicted in FIG. 9, is similar to the one depicted in FIG. 5, wherein it additionally comprises a recuperator (9), typically a counter-current heat exchanger located between the lowest-pressure compressor stage (1) and the second compressor stage (2) for preheating of air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also a closed-type heat rejection device (18) (typically a heat exchanger cooled by the environmental water or air). Assuming maximum cycle temperature of 1500 K (1227° C.), a recuperator effectiveness of 90%, an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜55.4% (three-pressures-fed radial GT) to ˜61.4% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜57.1% (three-pressures-fed radial GT) to ˜59.9% (four-pressures-fed radial GT).

The next similar preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine, depicted in FIG. 10, is similar to the one depicted in FIG. 5, wherein it additionally comprises a three-pressure-levels recuperator (9), typically a counter-current heat exchanger for preheating of air (or any other working gas) exiting each of the said compressor stages (multiple pressure levels) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a much greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system is can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″).and also a closed-type heat rejection device (18) (a heat exchanger cooled by the environmental water or air). In such a recuperated GT configuration, heat is being added to each of the three compressed working-gas streams coming from each of the said compressor stages after working-gas preheating in the said recuperator (9), using corresponding number of said isobaric or isochoric combustion chambers (4 a, 4 b, 4 c, . . . ) or said closed-type gas heaters (4 a″, 4 b″, 4 c″, . . . ). Assuming maximum cycle temperature of 1500 K (1227° C.), a recuperator effectiveness of 90%, an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same isentropic efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜61.3% (three-pressures-fed radial GT) to ˜64% (four-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜52.7% (three-pressures-fed radial GT) to ˜62.5% (five-pressures-fed radial GT).

The next preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 11, is similar to the one depicted in FIG. 5, wherein it additionally comprises said absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution and consisting of the same components already described in relation to the GT configuration depicted in FIG. 3. The refrigerant is being evaporated in the said evaporator (14) by pre-cooling of air (or any other working gas) at the inlet of the first stage (1) of the said multi-staged compressor, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18).

Another preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 12, is similar to the one depicted in FIG. 9 with said recuperator (9), wherein it additionally comprises said absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution and consisting of the same components already described in relation to the GT configuration depicted in FIGS. 3 and 11. The refrigerant is being evaporated in the said evaporator (14) by pre-cooling of air (or any other working gas) at the inlet of the first stage (1) of the said multi-staged compressor, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (−30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same efficiency for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜55.7% (three-pressures-fed radial GT) to ˜62% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields a theoretical cycle thermal efficiency ranging from ˜62.3% (two-pressures-fed radial GT) to ˜80.7% (five-pressures-fed radial GT).

Another preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 13, is similar to the one depicted in FIG. 12 with said three-pressure-levels recuperator (9) and said absorption refrigeration chiller (10), wherein said recuperator (9) is located between the lowest-pressure compressor stage (1) and the second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (−30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same for any of the compressor stages, with the compression pressure ratio (CPR) of 1.5:1 for each compressor stage, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜63.1% (three-pressures-fed radial GT) to ˜67.2% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields a theoretical cycle thermal efficiency ranging from ˜64.9% (two-pressures-fed radial GT) to ˜73.3% (four-pressures-fed radial GT).

Another preferred configuration of the open-cycle gas-turbine power-producing system using the rotating-plate radial gas turbine, depicted in FIG. 14, is similar to the one depicted in FIG. 5, wherein, instead of the said multi-staged compressor, said single-stage compressor (1) is being used associated with an axial turbine (5), or with an additionally provided second stage of the axial turbine (6), or, if need may be, with additionally provided multiple axial turbine stages, combined with the said compressor (1) in such a way as to feed said rotating-plate radial gas turbine (7) in cascades from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing. This way, the pressure of air (working gas) within a constant volume bordered by adjacent un-cooled rotating plates of the said radial gas turbine (7) and the said casing is being gradually increased, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. Said isobaric or isochoric combustion chamber (4) is located between the said compressor (1) and the first stage of the said axial gas turbine (5), adding heat to the total air (working gas) flow rate.

Related to the above configuration, FIG. 23 depicts corresponding thermodynamic temperature-entropy diagram of the open-cycle GT configuration of FIG. 14 with isobaric combustion chamber (max. cycle temperature 1500 K), using a three-pressures-fed rotating-plate radial gas turbine, two axial turbine stages and one compressor stage. Assuming an isentropic efficiency of the rotating-plate radial gas turbine of 90% and an isentropic efficiency of 85% of the compressor with the overall compression pressure ratio (CPR) of 30:1, this cycle configuration yields cycle thermal efficiency of ˜53.2%. It is to be noted that the real inlet temperature of the rotating-plate radial gas-turbine is somewhat lower than the maximum cycle temperature (1500 K), being ˜1328 K (1055° C.) as a result of mixing of all different-pressure-levels air-flow fractions between the rotating plates of the radial gas turbine (7). Assuming identical cycle input parameters and a five-pressures-fed rotating-plate radial gas turbine, the cycle configuration yields cycle thermal efficiency of ˜56.9%.

Again, a closed-cycle gas-turbine power-producing system similar to the multi-pressures-fed open-cycle radial gas-turbine power-producing system depicted in FIG. 14 is possible to configure and it is depicted in FIG. 15, wherein a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture is being circulated. Instead of said isobaric/isochoric combustion chamber (4), such a closed-cycle gas-turbine configuration would comprise a closed-type gas heater (4″), which can use any fuel source (heat input from: fossil fuels, nuclear fuels, renewable energy sources or waste heat), located between the outlet of the said compressor (1) and the inlet of the first stage of the said axial gas turbine (5). In addition, the configuration would have to include a closed-type heat rejection device (18), typically a heat exchanger cooled by the environmental water or air. Such a closed-cycle radial GT configuration would be also suitable for coal-fired and nuclear power plants of any type. For example, assuming helium as a working gas, maximum cycle temperature of 1123 K (850° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and an isentropic efficiency of 85% of the compressor with the overall compression pressure ratio (CPR) of 9:1, this cycle configuration yields cycle thermal efficiency ranging from ˜51.9% (three-pressures-fed radial GT) to ˜57.3% (five-pressures-fed radial GT).

The next preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine, depicted in FIG. 16, is similar to the one depicted in FIG. 14, wherein it additionally comprises said recuperator (9), typically a counter-current heat exchanger, located between the lowest-pressure compressor stage (1) and the second compressor stage (2) for preheating of air (or any other working gas) exiting the said compressor (1) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (−30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same efficiency for the compressor, with the overall compression pressure ratio (CPR) of 5:1, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜58.3% (two-pressures-fed radial GT) to ˜61.8% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜53.4% (three-pressures-fed radial GT) to ˜55.8% (five-pressures-fed radial GT).

Another preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine and a recuperator, depicted in FIG. 17, is similar to the one depicted in FIG. 16, wherein said recuperator (9) is located between the said compressor (1), being the lowest-pressure compressor stage, and additionally included second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a similar or greater cycle output and a somewhat higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (−30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same efficiency for both compressor stages, with the first-stage-compressor CPR of 4:1 and the second-stage-compressor CPR of 2:1, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜58.3% (two-pressures-fed radial GT) to ˜63.5% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜63.9% (one-pressure-fed radial GT) to ˜68% (five-pressures-fed radial GT).

The next preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine, depicted in FIG. 18, is similar to the one depicted in FIG. 14, wherein it additionally comprises said absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution and consisting of the same components already described in relation to the GT configuration depicted in FIGS. 3 and 11. The refrigerant is being evaporated in the said evaporator (14) by pre-cooling of air (or any other-working gas) at the inlet of the said compressor (1), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18).

Another preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine and said absorption refrigeration chiller (10), depicted in FIG. 19, is similar to the one depicted in FIG. 18, wherein it additionally comprises said recuperator (9), typically a counter-current heat exchanger for preheating of air (or any other working gas) exiting the said compressor (1) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (−30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same efficiency for the compressor, with the overall compression pressure ratio (CPR) of 5:1, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜62% (two-pressures-fed radial GT) to ˜65.2% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜63.4% (one-pressure-fed radial GT) to ˜68.2% (five-pressures-fed radial GT).

Finally, still another preferred configuration of the open-cycle gas-turbine power-producing system using a three-pressures-fed rotating-plate radial gas turbine, said absorption refrigeration chiller (10) and said recuperator (9), depicted in FIG. 20, is similar to the one depicted in FIG. 19, wherein said recuperator (9) is located between the said compressor (1), being the lowest-pressure compressor stage, and additionally included second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a similar or greater cycle output and a higher cycle thermal efficiency. A similar closed-cycle gas-turbine power-producing system can also be configured, using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture, and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18). Assuming maximum cycle temperature of 1500 K (1227° C.) in an open-cycle configuration, a recuperator effectiveness of 90%, minimum (refrigerated) cycle temperature of 243 K (˜30° C.), an isentropic efficiency of the rotating-plate radial gas turbine of 90% and the same efficiency for both compressor stages, with the first-stage-compressor CPR of 4:1 and the second-stage-compressor CPR of 2:1, this cycle configuration with isobaric heat addition yields cycle thermal efficiency ranging from ˜61.9% (two-pressures-fed radial GT) to ˜66.8% (five-pressures-fed radial GT). The same cycle configuration with isochoric heat addition yields cycle thermal efficiency ranging from ˜67.8% (one-pressure-fed radial GT) to ˜69.5% (five-pressures-fed radial GT).

All numbers expressing process or cycle parameters, cycle thermal efficiencies, specific cycle outputs, and so forth, used in this specification and claims are to be understood as being modified in all instances by the term “about” or “approximately”. The matter set forth in the foregoing description and accompanying drawings is offered by way of illustration only and not as a limitation. Since further modifications, applications or adaptations of the invention may become apparent to those skilled in the art, aim of the appended patent claims is to cover all such changes and modifications as fall within the true spirit and scope of the invention. 

1. A simple open-cycle gas-turbine power-producing system comprising a conventional compressor (1) of any suitable type (axial, radial, reciprocating, etc.) compressing air (working gas), an isobaric (constant-pressure) combustion chamber (4) fuelled by a gaseous or a liquid fuel, and a load, typically an electric generator (8), connected to said compressor (1) via a common rotating shaft, wherein the improvement comprises use of a “rotating-plate radial gas turbine” (7), which is essentially a rotating barrel with robust rectangular plates fitted into the turbine rotor (turbine blades) contained within a motionless rigid horizontal cylinder (casing), connected to the said compressor (1) and the said electric generator (8) via the same said common rotating shaft, fed with air (working gas) by the said compressor (1) and also exhausting the expanded combustion gas from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing, with an optional isochoric (constant-volume) heat addition (instead of said isobaric combustion chamber (4)) occurring in the spaces confined between adjacent un-cooled rotating plates and said static cylinder, where also a gaseous or a liquid fuel has to be injected and ignited by means of an electric spark, thus enabling a practical achievement of the Atkinson cycle (adiabatic compression and expansion, isochoric heat addition and isobaric heat rejection) with improved cycle thermal efficiency and greater cycle output.
 2. The gas-turbine power-producing system of the claim 1, wherein, in addition to the said compressor (1), the gas-turbine configuration comprises also a second stage of the compressor (2), accompanied by an intercooler (17) between the two said compressor stages, for achievement of a greater cycle output and a higher cycle thermal efficiency.
 3. The gas-turbine power-producing system of the claim 1, wherein the gas-turbine configuration additionally comprises an absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution, consisting of: a generator (11) heated by the exhaust gas from the said radial gas turbine (7) and evaporating the refrigerant from the solution, a condenser (12) for bringing the refrigerant into the liquid state, an absorber (16) for absorbing cold refrigerant vapor into the liquid solution, both said condenser (12) and said absorber (16) rejecting heat to cooling water or air from the environment, an expansion valve (13) for bringing the refrigerant to a lower temperature and an evaporator (14) for evaporating of the refrigerant by pre-cooling of air (working gas) at the inlet of the said compressor (1), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 4. The gas-turbine power-producing system of the claim 3, wherein, in addition to said compressor (1) and the corresponding said evaporator (14), the gas-turbine configuration comprises also a second stage of the compressor (2), accompanied by an intercooler (17) and a corresponding second-stage evaporator (15) for pre-cooling of air (working gas) at the inlet of the said second-stage compressor (2), both said intercooler (17) and said second-stage evaporator (15) located between the two said compressor stages, for achievement of a greater cycle output and a higher cycle thermal efficiency.
 5. The gas-turbine power-producing system of the claim 1, wherein, in addition to said compressor (1), the gas-turbine configuration comprises also a second stage of the compressor (2), a third stage of the compressor (3) and, if need may be, a plurality of the compressor stages, combined in such a way as to feed said rotating-plate radial gas turbine (7) in cascades from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing, thus gradually increasing pressure of air (working gas) within a constant volume bordered by adjacent un-cooled rotating plates of the said radial gas turbine (7) and the said casing, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency, whereas said isobaric/isochoric combustion chamber (4) is located between the highest-pressure stage of the said compressor and the said radial gas turbine (7).
 6. A closed-cycle gas-turbine power-producing system similar to the open-cycle gas-turbine power-producing system of the claim 5, wherein a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture is being circulated and wherein the gas-turbine configuration comprises a closed-type gas heater (4″) (heat input from: fossil fuels, nuclear fuels, renewable energy sources or waste heat) instead of said isobaric/isochoric combustion chamber (4), located between the outlet of the highest-pressure stage of said compressor and the inlet of said radial gas turbine (7), and also a closed-type heat rejection device (18), typically a heat exchanger cooled by the environmental water or air.
 7. The gas-turbine power-producing system of the claim 5, wherein the gas-turbine configuration does not comprise said isobaric/isochoric combustion chamber (4) or any other heat addition at all and wherein it employs solely the effect of a gradual increase of the air (working-gas) pressure within a constant volume bordered by adjacent rotating plates of said radial gas turbine (7) and the said static casing as a result of feeding the said rotating-plate radial gas turbine (7) in cascades from the radial direction by combining multiple stages of the said compressor, thus achieving a positive cycle output and the cycle thermal efficiency of 100%.
 8. A closed-cycle gas-turbine power-producing system similar to that of the claim 6, wherein the said closed-type gas heater (4″) is located: (a) either between the outlet of the highest-pressure stage of said compressor and the inlet of said radial gas turbine (7) (for higher-temperature heat addition), or (b) between the outlet of the said radial gas turbine (7) and the inlet of the lowest-pressure stage of the said compressor (for lower-temperature heat addition), whereas the said closed-cycle gas-turbine configuration does not comprise any heat rejection device, thus achieving a positive cycle output and the cycle thermal efficiency of 100%.
 9. The open-cycle gas-turbine power-producing system of the claim 5, or a similar closed-cycle gas-turbine power-producing system using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18) (typically a heat exchanger cooled by environmental water or air), wherein said open-cycle or said closed-cycle gas-turbine configuration additionally comprises an absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution, consisting of: a generator (11) heated by exhaust gas from the said radial gas turbine (7) and evaporating the refrigerant from the solution, a condenser (12) for bringing the refrigerant into the liquid state, an absorber (16) for absorbing cold refrigerant vapor into the liquid solution, both said condenser (12) and said absorber (16) rejecting heat to cooling water or air from the environment, an expansion valve (13) for bringing the refrigerant to a lower temperature and an evaporator (14) for evaporating of the refrigerant by pre-cooling of air (or any other working gas) at the inlet of the first stage (1) of the said multi-staged compressor, enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 10. The open-cycle gas-turbine power-producing system of the claim 5, or a similar closed-cycle gas-turbine power-producing system using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18) (typically a heat exchanger cooled by the environmental water or air), wherein said open-cycle or said closed-cycle gas-turbine configuration additionally comprises a recuperator (9), which is typically a counter-current heat exchanger for preheating of air (or any other working gas) exiting each of the said compressor stages by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency, whereas heat is being added to each of the compressed working-gas streams coming from each of the said compressor stages after said working-gas preheating in the said recuperator (9), using corresponding number of said combustion chambers (4 a, 4 b, 4 c, . . . ) or said closed-type gas heaters (4 a″, 4 b″, 4 c″, . . . ).
 11. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 10, wherein said recuperator (9) is located between the lowest-pressure compressor stage (1) and the second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a lower cycle output and a similar cycle thermal efficiency.
 12. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 10, wherein said open-cycle or said closed-cycle gas-turbine configuration with said recuperator (9) additionally comprises an absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution, consisting of: a generator (11) heated by exhaust gas from the said radial gas turbine (7) and evaporating the refrigerant from the solution, a condenser (12) for bringing the refrigerant into the liquid state, an absorber (16) for absorbing cold refrigerant vapor into the liquid solution, both said condenser (12) and said absorber (16) rejecting heat to cooling water or air from the environment, an expansion valve (13) for bringing the refrigerant to a lower temperature and an evaporator (14) for evaporating of the refrigerant by pre-cooling of air (or any other working gas) at the inlet of the first stage (1) of the said multi-staged compressor, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 13. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 12, wherein said recuperator (9) is located between the lowest-pressure compressor stage (1) and the second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a lower cycle output and a similar cycle thermal efficiency.
 14. An open-cycle gas-turbine power-producing system similar to the gas-turbine power-producing system of the claim 5, wherein, instead of the said multi-staged compressor, said single-stage compressor (1) is being used coupled with an axial turbine (5), or with an additionally provided second stage of the axial turbine (6), or, if need may be, with additionally provided multiple axial turbine stages, combined with the said compressor (1) in such a way as to feed said rotating-plate radial gas turbine (7) in cascades from the radial direction thru corresponding side/top/bottom openings in the said cylinder/casing, thus gradually increasing pressure of air (working gas) within a constant volume bordered by adjacent un-cooled rotating plates of the said radial gas turbine (7) and the said casing, thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency, whereas said isobaric/isochoric combustion chamber (4) is located between the said compressor (1) and the first stage of the said axial gas turbine (5), adding heat to the total air (working gas) flow rate.
 15. A closed-cycle gas-turbine power-producing system similar to the open-cycle gas-turbine power-producing system of the claim 14, wherein a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture is being circulated and wherein the gas-turbine configuration comprises a closed-type gas heater (4″) (heat input from: fossil fuels, nuclear fuels, renewable energy sources or waste heat) instead of said isobaric/isochoric combustion chamber (4), located between the outlet of the said compressor (1) and the inlet of the first stage of the said axial gas turbine (5), and also a closed-type heat rejection device (18), typically a heat exchanger cooled by the environmental water or air.
 16. The open-cycle gas-turbine power-producing system of the claim 14, or a similar closed-cycle gas-turbine power-producing system using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18), wherein said open-cycle or said closed-cycle gas-turbine configuration additionally comprises a recuperator (9), which is typically a counter-current heat exchanger for preheating of air (or any other working gas) exiting the said compressor (1) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 17. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 16, wherein said recuperator (9) is located between the said compressor (1), being the lowest-pressure compressor stage, and additionally included second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a similar or greater cycle output and a higher cycle thermal efficiency.
 18. The open-cycle gas-turbine power-producing system of the claim 14, or a similar closed-cycle gas-turbine power-producing system using a suitable working gas (helium, CO₂, nitrogen, etc.) or a gas mixture and comprising said closed-type gas heater (4″) and also said closed-type heat rejection device (18), wherein said open-cycle or said closed-cycle gas-turbine configuration additionally comprises an absorption refrigeration chiller (10), utilizing an appropriate refrigerant-carrier (ammonia-water or water-lithium bromide) mixture/solution, consisting of: a generator (11) heated by exhaust gas from the said radial gas turbine (7) and evaporating the refrigerant from the solution, a condenser (12) for bringing the refrigerant into the liquid state, an absorber (16) for absorbing cold refrigerant vapor into the liquid solution, both said condenser (12) and said absorber (16) rejecting heat to cooling water or air from the environment, an expansion valve (13) for bringing the refrigerant to a lower temperature and an evaporator (14) for evaporating of the refrigerant by pre-cooling of air (or any other working gas) at the inlet of the said compressor (1), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 19. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 18, wherein said open-cycle or said closed-cycle gas-turbine configuration with said absorption refrigeration chiller (10) additionally comprises a recuperator (9), typically a counter-current heat exchanger for preheating of air (or any other working gas) exiting the said compressor (1) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a greater cycle output and a higher cycle thermal efficiency.
 20. The open-cycle or the closed-cycle gas-turbine power-producing system of the claim 19, wherein said recuperator (9) is located between the said compressor (1), being the lowest-pressure compressor stage, and additionally included second compressor stage (2), preheating air (or any other working gas) at the outlet of the said lowest-pressure compressor stage (1) (prior to the inlet of the said second compressor stage (2)) by exhaust gas from the said radial gas turbine (7), thus enabling achievement of a similar or greater cycle output and a higher cycle thermal efficiency. 